EVAPORATOR HEAT EXCHANGER FOR PREVENTING ICE BUILD-UP

Abstract

A system includes a compressor for increasing the pressure of a refrigerant; a condenser heat exchanger arranged downstream of the compressor for receiving a high pressure refrigerant output from the compressor and for transferring heat from the high pressure refrigerant to a source of water; an expansion device for reducing the pressure of a refrigerant; and an evaporator heat exchanger for extracting heat from ambient air. The evaporator heat exchanger includes a first tube bank having a first inlet arranged to receive a high pressure refrigerant output from the condenser and a first outlet fluidly coupled to the expansion device; and a second tube bank having a second inlet arranged to receive a low pressure refrigerant output from the expansion device, and a second outlet fluidly coupled to an inlet of the compressor.

Description

FOREIGN PRIORITY

This application claims priority to European Patent Application No. 21203005.0, filed Oct. 15, 2021, and all the benefits accruing therefrom under 35 U.S.C. § 119, the contents of which in its entirety are herein incorporated by reference.

TECHNICAL FIELD OF INVENTION

The present invention relates to an air-to-water heat pump system for preventing and/or limiting the build-up of ice on an evaporator heat exchanger, and a method of preventing and/or limiting the build-up of ice on an evaporator heat exchanger in an air-to-water heat pump.

BACKGROUND OF THE INVENTION

In recent years, heat pumps have been increasing in popularity, with around 100,000 heat pump units being installed each year. This rise in popularity can be attributed to the fact that heat pumps are now considered one of the most efficient, cost-effective and environmentally friendly methods of heating. Heat pump systems can be used to provide heat to homes, offices, and other indoor spaces by transferring thermal energy from one point to another. Two of the most common types of heat pumps are air-source heat pumps and ground-source heat pumps; these draw heat from outdoor air or the ground, respectively, using an evaporator heat exchanger, and transfer it to an indoor space via a refrigerant, where it is released using a condenser heat exchanger. Though they require a small amount of electricity to run, heat pumps are considered to be “clean”, as they do not rely upon the burning of fuel to produce heat.

As heat pump systems increase in popularity across the world, issues have begun to arise as they are used more frequently in colder climates. This is especially the case for air-source heat pumps, which decrease in efficiency during the colder months when there is less heat to extract from the outside air. When the outdoor ambient temperature is low, it becomes possible for ice to form on the evaporator heat exchanger. This is because water vapour present in the air will be cooled as it passes over the evaporator heat exchanger, causing it to condense and, if cold enough, freeze. If enough ice forms, it may cover a substantial portion of the outer surface of the evaporator heat exchanger, thus preventing the air from contacting the evaporator heat exchanger and hindering the exchange of heat between the air and the refrigerant circulating within the evaporator heat exchanger.

There are a number of existing solutions to address this problem of ice formation. For example, some heat pumps include electrical heaters that are positioned close to the evaporator heat exchanger and are turned on at certain time intervals in order to melt any ice that may have built up. Alternatively, some heat pumps can be placed in a “defrost mode”, wherein the circulation of refrigerant through the heat pump is reversed, causing the evaporator heat exchanger to temporarily become a condenser heat exchanger and vice versa. As the evaporator heat exchanger is now a condenser heat exchanger, it expels heat and causes ice that has built up on the external surface of the heat exchanger to melt. However, both of these options increase the complexity, energy needs, footprint, and cost of the heat pump system.

SUMMARY OF THE INVENTION

According to a first aspect of the invention, an air-to-water heat pump system is provided comprising: a compressor for increasing the pressure of a refrigerant; a condenser heat exchanger arranged downstream of the compressor for receiving a high pressure refrigerant output from the compressor and for transferring heat from the high pressure refrigerant to a source of water; an expansion device for reducing the pressure of a refrigerant; and an evaporator heat exchanger for extracting heat from ambient air, the evaporator heat exchanger comprising: a first tube bank having a first inlet arranged to receive a high pressure refrigerant output from the condenser and a first outlet fluidly coupled to the expansion device; and a second tube bank having a second inlet arranged to receive a low pressure refrigerant output from the expansion device, and a second outlet fluidly coupled to an inlet of the compressor; wherein the first and second tube banks are arranged in close proximity to one another such that, in use, heat from the high pressure refrigerant passing through the first tube bank is transferred to the second tube bank in order to limit and/or prevent ice build-up on an external surface of the second tube bank.

The heat pump system of the first aspect therefore utilises the heat from the high pressure, high temperature refrigerant in the first tube bank in order to provide heat to the outer surface of the second tube bank, which contains low pressure, low temperature refrigerant. This is achieved by placing the two tube banks in close proximity to one another, thus enabling heat to be transferred between the two. In doing this, the temperature of the outer surface of the second tube bank can be raised by a sufficient amount to ensure that water vapour present in the air does not freeze on the outer surface of the second tube bank. As a result, any drops in efficiency that may arise due to excessive ice formation are prevented. Prior art systems typically rely on an external heat source (such as an electric heater) or a “defrost mode” of operation, which are both relatively energy intensive, to achieve this effect. In contrast, the present invention utilises heat already within the system to prevent ice formation, thus providing a simple, compact and more energy efficient solution to the problem of ice formation.

It will be appreciated that the evaporator heat exchanger comprises the first and second tube banks in a single, integrated unit. Hence, the first and second tube banks may be arranged within the same, single housing. This may be similar to typical prior art evaporator heat exchangers that comprise multiple rows. However, in the present invention, one of the rows is fluidly connected to the high pressure side of the system for receiving refrigerant before the refrigerant has been passed through the expansion device.

The heat pump system comprises a compressor for increasing the pressure of the refrigerant. The refrigerant may be, for example, R-32 refrigerant or propane. The compressor may receive a low pressure refrigerant from, for example, the evaporator heat exchanger, and may output a high pressure refrigerant. The compressor may be configured to output a gas phase refrigerant. The compressor may be a scroll compressor. The scroll compressor may comprise two spiral-shaped scrolls. A first spiral-shaped scroll of the two spiral-shaped scrolls may be configured to remain stationary whilst a second spiral-shaped scroll of the two spiral-shaped scrolls may be configured to orbit around the first spiral-shaped scroll, such that the refrigerant travelling between the spiral-shaped scrolls is forced into increasingly smaller areas and compressed.

The condenser heat exchanger is configured to receive the high pressure refrigerant output from the compressor. The condenser heat exchanger may be configured such that a source of water passes through the condenser heat exchanger and exchanges heat with the high pressure refrigerant. The condenser heat exchanger may be configured such that heat from the high pressure refrigerant is transferred to the source of water, thereby heating the water and cooling the refrigerant. The source of water may then pass to, for example, a radiator within an indoor space. The condenser heat exchanger may be considered to be a water-cooled compressor heat exchanger. The condenser heat exchanger may comprise a condenser tube bank through which the water flows. The condenser heat exchanger may be configured such that the high pressure refrigerant contacts the condenser tube bank and exchanges heat with the condenser tube bank, and the water within. The condenser heat exchanger may be configured to condense at least a portion of the high pressure refrigerant, causing that portion to change phase from a gaseous refrigerant to a liquid refrigerant.

The first tube bank of the evaporator heat exchanger comprises a first inlet configured to receive the high pressure refrigerant that is output from the condenser. Typical prior art heat pump systems may be configured to directly pass this high pressure refrigerant into an expansion device, where it may be reduced in pressure, which may cause at least a portion of the liquid refrigerant to change phase to a gaseous refrigerant. This may also lead to cooling of the refrigerant. The heat pump system of the prior art device may then be configured to pass the resultant low pressure refrigerant to an evaporator heat exchanger, where it may be evaporated and then fed back to the compressor. In the present system, the heat present in the high pressure refrigerant is utilised in order to limit and/or prevent ice build-up on the external surface of the second tube bank. As such, the first inlet of the first tube bank is configured to receive the high pressure refrigerant from the condenser heat exchanger prior to the expansion device receiving the refrigerant. In the first tube bank, heat from the high pressure refrigerant is passed from the first tube bank to the second tube bank.

The first outlet of the first tube bank is in fluid communication with the expansion device. The expansion device is configured to receive the high pressure fluid from the first outlet. The expansion device may be configured to reduce the pressure of the refrigerant such that it becomes a low pressure refrigerant. This reduction in pressure preferably causes at least a portion of the refrigerant to change state from a liquid to a gas. This may also lower the temperature of the refrigerant. The expansion device may be an electronic expansion valve.

The second inlet of the second tube bank of the evaporator heat exchanger is configured to receive the low pressure refrigerant output from the expansion device. The low pressure refrigerant in the second tube bank may have a lower temperature than the high pressure refrigerant in the first tube bank. As a result, the outer wall temperature of the first tube bank may be higher than the outer wall temperature of the second tube bank. The heat pump system may be configured such that heat is released from the high pressure refrigerant in the first tube bank and is passed to the low pressure refrigerant in the second tube bank. The heat pump system may be configured such that heat from the high pressure refrigerant in the first tube bank is absorbed by the low pressure refrigerant in the second tube bank. This transfer of heat may cause the outer wall temperature of the second tube bank to be sufficiently high to limit and/or prevent ice formation. The heat pump system may be configured, in use, to cause refrigerant to flow through the first tube bank in a direction that is counter-current to the flow of refrigerant through the second tube bank. This may result in improved heat transfer between the first and second tube banks and/or the refrigerant flowing through the first and second tube banks.

As used herein the term “close proximity” is used to refer to an arrangement in which the first and second tube banks are positioned close to one another so as to permit heat to be transferred from refrigerant flowing through the first tube bank to the second tube bank. A gap may exist between the first tube bank and the second tube bank. The gap may be small in size in order to permit heat transfer from the refrigerant in the first tube bank to the second tube bank. The size of the gap may vary across the length of the first and second tube banks, but may always be 1 mm or less, or preferably 0.5 mm or less. Alternatively, or in addition, the average size of the gap across the length of the first and second tube banks may be 1 mm or less, or optionally 0.5 mm or less. The gap may be sized such that water droplets, for example water droplets that condense from the ambient air, may bridge the gap between the first and second tube banks to enable heat to be transferred between the first and second tube banks via conduction via the water droplets. The first tube bank may be in direct contact with the second tube bank at points along its length, such that the size of the gap at those points is 0 mm. Additionally or alternatively, the first tube bank may be in contact with the second tube bank along its entire length. This close proximity between the first tube bank and the second tube bank may allow the first tube bank and the second tube bank to be in a conductive heat exchange relationship with one another. In other words, the heat pump system may be configured such that, in use, heat is transferred between the first tube bank and the second tube bank through conduction.

The heat pump system may also be configured such that, in use, heat is transferred between the first tube bank and the second tube bank through other means of heat transfer, such as radiation and/or convection.

The first tube bank may be arranged in a heat exchange relationship with ambient air such that, in use, heat is transferred from the high pressure refrigerant to the ambient air. The high pressure refrigerant therefore may be cooled by the ambient air, and the ambient air may be warmed by the high pressure refrigerant. It will be appreciated that this cooling of the high pressure refrigerant may be in addition to cooling of the high pressure refrigerant caused as a result of heat transfer from the high pressure refrigerant to the second tube bank. The first tube bank may be configured to provide subcooling to at least a portion of the high pressure refrigerant. That is, the first tube bank may be arranged such that, in use, at least a portion of the high pressure refrigerant undergoes cooling, or subcooling, as it passes through the first tube bank. This subcooling may be achieved through heat exchange with the low pressure refrigerant in the second tube bank and/or the ambient air. The first tube bank may be configured to provide, in use, up to 20° C., for instance from 0° C. to 20° C., of subcooling to the high pressure refrigerant.

The first tube bank may be configured such that, in use, any remaining gaseous refrigerant in the high pressure refrigerant is condensed and changed to the liquid phase. The first tube bank may therefore be configured to remove any flash gas (i.e. gaseous refrigerant) from the refrigerant prior to the expansion device. This may result in the expansion device receiving a larger amount of refrigerant than if the refrigerant were not subcooled, thus increasing the heating capacity of the system.

The second tube bank may be arranged in a heat exchange relationship with ambient air such that, in use, heat is transferred from the ambient air to the low pressure refrigerant. The low pressure refrigerant may therefore be heated by the ambient air, and the ambient air may therefore be cooled by the high pressure refrigerant. The second tube bank may be arranged to provide superheating to at least a portion of the low pressure refrigerant. That is, the second tube bank may be arranged such that, in use, at least a portion of the low pressure refrigerant undergoes heating, or superheating, as it passes through the second tube bank. This superheating may be achieved through heat exchange with the high pressure refrigerant and/or the ambient air. The second tube bank may be configured to provide, in use, up to 20° C., e.g. between 1° C. and 20° C., of superheating to the low pressure refrigerant.

The heat pump system may be configured such that, in use, at least a portion of the low pressure refrigerant passing through the second tube bank is evaporated, thereby causing that portion to change phase from a liquid refrigerant to a gaseous refrigerant. Preferably, the heat pump system is configured such that any liquid low pressure refrigerant is vaporised as it passes through the second tube bank, such that the entire low pressure refrigerant is in the gas phase when it leaves the second outlet. The second tube bank may therefore ensure that no, or very little, liquid refrigerant is fed to the compressor. This may be advantageous, as the presence of liquid refrigerant in the compressor can hinder its performance by imposing unwanted hydraulic forces and/or “washing away” the protective and lubricating effects of any compressor oil present in the compressor.

The first tube bank may comprise one or more rows, for instance, two, three, four or more rows. The one or more rows of the first tube bank may be arranged in parallel to one another. The one or more rows of the first tube bank may be fluidly coupled to one another in series. The first inlet may be configured to feed refrigerant into a first row of the one or more rows of the first tube bank, and the first outlet may be configured to receive refrigerant from a last row of the one or more rows of the first tube bank. The first tube bank may be configured, in use, to transport refrigerant sequentially through the rows. For example, the first tube bank may comprise three rows. The first inlet may be configured to feed refrigerant into the first row of the first tube bank, and the first outlet may be configured to receive refrigerant from the third row of the first tube bank. The first tube bank may be configured to transport the refrigerant firstly through the first row of the first tube bank, then through the second row of the first tube bank, and finally through the third row of the first tube bank.

One or more or all of the rows of the one or more rows of the first tube bank may comprise one or more first tubes that are connected in series. The one or more first tubes may be connected so as to form a serpentine shape. Alternatively, the one or more first tubes may be connected in a coil shape. This arrangement of serpentine or coiled first tubes being arranged parallel to one another in rows may increase the surface area of the first tube bank, thus improving its heat transfer capabilities.

The first tube bank may comprise multiple, i.e. more than one, refrigerant circuits, for instance two, three or more circuits. In other words, the first tube bank may be divided into two or more circuits. The first inlet may be connected to a first refrigerant distributor downstream of the first inlet. The first refrigerant distributor may be configured to feed the refrigerant into the one or more refrigerant circuits, such that the refrigerant is distributed, preferably equally, across the refrigerant circuits. The first outlet may be connected to a first refrigerant header upstream of the first outlet. The first refrigerant header may be configured to receive refrigerant from each of the one or more refrigerant circuits, such that the refrigerant from each of the refrigerant circuits is re-combined into a single flow of refrigerant. The circuits may flow through each of the one or more rows of the first tube bank. That is, each of the rows of the tube bank may comprise multiple circuits, with a circuit in one of the rows being connected in series with a circuit in one or more of the other rows. The first tube bank may be arranged such that the refrigerant flowing through the circuits is re-combined only after it has passed through one or more circuits in each row. The use of multiple refrigerant circuits introduces redundancy into the first tube bank, and the heat pump system as a whole. The provision of multiple refrigerant circuits may result in a lower pressure drop in the high pressure refrigerant, and may increase the thermal efficiency of the first tube bank. In this way, the provision of multiple refrigerant circuits may optimise the heat exchange performance of the first tube bank.

The second tube bank may comprise one or more rows, for instance, two three, four or more rows. The one or more rows of the second tube bank may be arranged in parallel to one another. The one or more rows of the second tube bank may be fluidly coupled to one another in series. The second inlet may be configured to feed refrigerant into a first row of the one or more rows of the second tube bank, and the second outlet may be configured to receive refrigerant from the last row of the one or more rows of the second tube bank. The second tube bank may be configured, in use, to transport refrigerant sequentially through the rows. For example, the second tube bank may comprise three rows. The second inlet may be configured to feed refrigerant into the first row of the second tube bank, and the second outlet may be configured to receive refrigerant from the third row of the second tube bank. The second tube bank may be configured to transport the refrigerant firstly through the first row of the second tube bank, then through the second row of the second tube bank, and finally through the third row of the second tube bank. The first or last row of the one or more rows of the second tube bank may be adjacent to the first tube bank, for instance the first or last row of the one or more rows of the first tube bank. The one or more rows of the second tube bank may be arranged parallel to the first tube bank, for instance parallel to the one or more rows of the first tube bank.

One or more or all of the rows of the one or more rows of the second tube bank may comprise one or more second tubes that are connected in series. The one or more second tubes may be connected in a serpentine shape. Alternatively, the one or more second tubes may be connected in a coil shape. This arrangement of serpentine or coiled second tubes being arranged parallel to one another in rows may increase the surface area of the second tube bank, thus improving its heat transfer capabilities. The second tubes may extend in a direction parallel to the direction in with the first tubes extend.

The second tube bank may comprise one or more refrigerant circuits, for instance two, three or more circuits. In other words, the second tube bank may be divided into two or more circuits. The second inlet may be connected to a second refrigerant distributor downstream of the second inlet. The second refrigerant distributor may be configured to feed the refrigerant into the one or more refrigerant circuits, such that the refrigerant is distributed, preferably equally, across the refrigerant circuits. The second outlet may be connected to a second refrigerant header upstream of the second outlet. The second refrigerant header may be configured to receive refrigerant from each of the one or more refrigerant circuits, such that the refrigerant from each of the refrigerant circuits is re-combined into a single flow of refrigerant. The circuits may flow through each of the one or more rows of the second tube bank. That is, each of the rows of the second tube bank may comprise multiple circuits, with a circuit in one of the rows being connected in series with a circuit in one or more of the other rows. The second tube bank may be arranged such that the refrigerant flowing through the circuits is re-combined only after it has passed through one or more circuits in each row. The use of multiple refrigerant circuits introduces redundancy into the second tube bank, and to the heat pump system as a whole. The provision of multiple refrigerant circuits may result in a lower pressure drop in the low pressure refrigerant, and may increase the thermal efficiency of the second tube bank. In this way, the provision of multiple refrigerant circuits may optimise the heat exchange performance of the second tube bank.

The first and/or second tube bank may comprise a round-tube and plate-fin (RTPF) heat exchanger. Optionally, the first and/or second tube bank may comprise a micro-channel heat exchanger (MCHE). The first and/or second tubes of the first and/or second tube bank may comprise copper and/or aluminium tubes.

The first tube bank may comprise a first plurality of fins. The first plurality of fins may be in direct contact with the first tubes of the first tube bank. The first plurality of fins may be a first plurality of plate fins. The first tubes of the first tube bank may extend through the first plurality of plate fins, such that the first plurality of plate fins extend in a direction (substantially) perpendicular to the longitudinal direction of the first tubes, and such that the plate fins extend (substantially) perpendicular to the flow of refrigerant through the first tubes. The first plurality of fins may comprise copper and/or aluminium fins.

The second tube bank may comprise a second plurality of fins. The second plurality of fins may be in direct contact with the second tubes of the second tube bank. The second plurality of fins may be a second plurality of plate fins. The second tubes of the second tube bank may extend through the second plurality of plate fins, such that the second plurality of plate fins extend in a direction (substantially) perpendicular to the longitudinal direction of the second tubes, and such that the plate fins extend (substantially) perpendicular to the flow of refrigerant through the second tubes. The second plurality of fins may comprise copper and/or aluminium fins.

The first plurality of fins may extend in the same direction as the second plurality of fins. At least some of the first plurality of fins may be in close proximity to at least some of the second plurality of fins, such that, in use, heat from the high pressure refrigerant passing through the first tube bank is transferred to the second tube bank via the first plurality of fins and the second plurality of fins in order to limit and/or prevent ice build-up on the external surface of the second tube bank. At least a portion of the first plurality of fins may be in direct contact with at least a portion of the second plurality of fins. Alternatively, there may be a gap between the first plurality of fins and the second plurality of fins. The gap may vary across the length of the first and second tube banks, but may always have a size of 1 mm or less, preferably 0.5 mm or less. Alternatively, or in addition, the average size of the gap across the length of the first and second tube banks may be 1 mm or less, or optionally 0.5 mm or less.

The first and/or second plurality of fins may be arranged in a heat exchange relationship with the ambient air, such that, in use, heat is transferred between the first and/or second plurality of fins and the ambient air. The first and/or second plurality of fins may increase the surface area of the first and/or second tube banks, and may therefore improve their heat exchange capabilities.

The heat pump system may be configured such that the temperature of all or at least a portion of the external surface of the first tube bank, when in use, is at least 50° C. In particular, the external temperature of the row of the first tube bank that is closest to the second tube bank may be at least 50° C. In this way, it may be ensured that the first tube bank is sufficiently hot for heating the second tube bank to prevent ice build-up thereon. The heat pump system may be configured such that the temperature of the external surface of the second tube bank, when in use, is greater than 0° C. Particularly, the temperature of the external surface of the row of the second tube bank that is furthest from the first tube bank may be greater than 0° C. In this way, it may be ensured that ice build-up on the entirety of the second tube bank is limited and/or prevented. It will be appreciated that, in use, the external surface temperature of the rows of the second tube bank may decrease the further away the row is from the first tube bank. However, the external surface temperature of all of the rows of the second tube bank may be sufficiently high to prevent and/or limit ice build-up thereon.

The term “freeze limit” is used herein to refer to the minimum ambient temperature of the air at which ice will not form on the evaporator heat exchanger. The freeze limit of prior art heat pump systems may typically be 10° C. or greater. This presents problems in countries with cold climates, where temperatures frequently fall below 10° C. in the winter. The heat pump system of the present invention may be configured such that it has a freeze limit of less than 10° C., and preferably 7° C. or less. This lowering of the freeze limit provides significant benefits to the heat pump system, as it may prevent and/or limit ice build-up at much lower ambient temperatures than the prior art heat pump systems. The low freeze limit also may ensure that the rate of ice build-up at a given temperature is reduced. As such, an excessive build-up of ice, and the disadvantages associated with such an excessive build-up of ice, may be avoided. For example, at ambient temperatures at or above the freeze limit, the heat pump system of the present invention may not ever be required to enter a “defrost” mode. This may be advantageous, as such a mode typically requires large amounts of energy and disruption to the system. In addition, the prevention and/or limitation of ice build-up may lead to improvements in efficiency of the heat pump system. In other words, the coefficient of performance of the heat pump system may be significantly improved. The heat pump system of the present invention may, for example, be up to 20% more efficient than heat pump systems of the prior art. Moreover, the heat pump system of the present invention may be used in wider range of operating conditions, and at lower ambient temperatures.

The heat pump system may comprise a fan for creating a flow of air over the evaporator heat exchanger. The fan, first tube bank and the second tube bank may be arranged such that, in use, a flow of air created by the fan first passes over the first tube bank and then passes over the second tube bank. Alternatively, the fan, first tube bank and the second tube bank may be arranged such that, in use, a flow of air generated by the fan first passes over the second tube bank and then passes over the first tube bank. In the case where the fan is configured to create a flow of air that first passes over the first tube bank, the air flow can be heated through heat exchange with high pressure refrigerant flowing through the first tube bank prior to it passing over the second tube bank; consequently, the air passing over the second tube bank may provide heat to the second tube bank such that the external surface temperature of the second tube bank is raised. The fan may be arranged adjacent to the first tube bank, such that it is closer to the first tube bank than the second tube bank. Alternatively, the fan may be arranged adjacent to the second tube bank, such that it is closer to the second tube bank than the first tube bank. The arrangement of the fan may depend on the desired flow direction of the air. The fan may be configured to provide a flow of air that is counter-current to the flow of refrigerant through the first and/or second tube bank. This may result in improved heat transfer between the air and the first and/or second tube bank.

The heat pump system may be operable in a heating mode and a cooling mode. For this purpose, the heat pump system may include a reversing valve, for instance a four-way valve, for reversing the flow of refrigerant through the system. In a heating mode, the system may be arranged such that refrigerant output from the compressor is passed, sequentially, to the condenser heat exchanger, to the first tube bank of the evaporator heat exchanger, to the expansion device, to the second tube bank of the evaporator heat exchanger and back to an inlet of the compressor. In the cooling mode, the system may be arranged such that refrigerant output from the compressor is passed, sequentially, to the second tube bank, the expansion device, the first tube bank, the condenser heat exchanger and back to an inlet of the compressor. In the cooling mode the evaporator heat exchanger may operate as a condenser heat exchanger for expelling heat to the air. In this way, the cooling mode may function as a “defrost” mode, enabling ice that has built up on the evaporator heat exchanger to be melted. Hence, the terms “cooling mode” and “defrost mode” may be used interchangeably herein.

Conventional reversible heat pump systems may include a receiver, i.e. a vessel for storing liquid refrigerant, due to the difference in volume between evaporator heat exchangers and condenser heat exchangers. Evaporator heat exchangers typically have a greater volume than condenser heat exchangers, and hence generally have a greater capacity than condenser heat exchangers. As such, switching a system from a heating mode to a cooling mode, i.e. when an evaporator heat exchanger is operated as a condenser heat exchanger, may result in excess liquid refrigerant within the system that could reach the compressor without being evaporated. This could cause damage to the compressor and/or affect system performance. In order to account for this, reversible heat pump systems may typically include a receiver arranged in the refrigerant flow path, e.g. between the condenser heat exchanger and the evaporator heat exchanger, to intercept the liquid refrigerant and retain it to prevent it from reaching the compressor. However, the heat pump system of the present invention may not require a receiver to account for the excess liquid refrigerant present in in the system during a cooling mode. Rather, the first tube bank may address the issue of excess liquid refrigerant by providing a means to store the excess liquid refrigerant in the cooling mode. That is, the first tube bank may function similar to a receiver to retain excess liquid refrigerant in the system during the cooling mode. As such, there may be no need for a receiver within the heat pump system, and the overall complexity of the heat pump system may be reduced.

The heat pump system may comprise a controller. The controller may be configured to control the operation of the various components within the heat pump system. For example, the controller may be configured to control the degree of opening of the electronic expansion valve. The controller may be configured to control the system automatically, for example in response to data measured by certain sensors within the heat pump system, and/or may be configured to control the system in response to certain inputs from an operator. The controller may be configured to control the operation of the heat pump system such that the heat pump system has a desired heating capacity. The controller may be configured to switch the mode of the heat pump system when desired, for instance from a heating mode to a defrost mode. For this purpose, the controller may be configured to control operation of the four-way valve.

According to a second aspect of the invention, a method of preventing and/or limiting the build-up of ice on an evaporator heat exchanger in situ in an air-to-water heat pump system is provided. The method comprises: passing a high pressure refrigerant through a first tube bank of the evaporator heat exchanger in order to cool the high pressure refrigerant though heat exchange between the high pressure refrigerant and ambient air; reducing the pressure of the cooled high pressure refrigerant, so as to provide a low pressure refrigerant; and passing the low pressure refrigerant through a second tube bank of the evaporator heat exchanger in order to warm the low pressure refrigerant through heat exchange between the low pressure refrigerant and ambient air; wherein the first and second tube banks are arranged in close proximity to one another such that heat from the high pressure refrigerant passing through the first tube bank is transferred to the second tube bank in order to heat at least a portion of an external surface of the second tube bank, thereby limiting and/or preventing ice from building up on the external surface of the second tube bank.

The method may be carried out using the air-to-water heat pump system described above in respect of the first aspect and may include any one or more or all of the optional features described above. Hence, a gap of 1 mm or less, or 0.5 mm or less may be present between the first and second tube banks. Alternatively, or in addition, the first tube bank may be in direct contact with the second tube bank at points along its length. The controller of the first aspect may be configured to control components of the air-to-water heat pump system, such as the compressor, the reversing valve, and/or the expansion device, in order to perform the method of the second aspect.

The ambient air may be cooled though heat exchange with the low pressure refrigerant to such an extent that water vapour present within the air may condense. This may result in water droplets forming on the external surface(s) of the first and/or second tube bank. The water droplets may bridge the gap between the first and second tube banks, i.e. one or more water droplets may be in contact with both the first and second tube banks. This may allow heat to be transferred between the first and second tube banks via conduction via the water droplet(s). This may improve the ability of the system to heat the second tube bank via the high pressure refrigerant within the first tube bank.

The method may comprise increasing the pressure of the refrigerant with a compressor, which may be a scroll compressor. The method may include providing the high pressure refrigerant using the compressor. The method may include passing the high pressure refrigerant from the compressor to a condenser heat exchanger.

The method may comprise using a condenser heat exchanger to cool the high pressure refrigerant prior to passing the refrigerant to the first tube bank. The method may comprise cooling the high pressure refrigerant using the condenser heat exchanger through heat exchange between the high pressure refrigerant and a water source. The method may include flowing the water source through tubes of the condenser heat exchanger, and may comprise warming the water source through heat exchange with the high pressure refrigerant. The method may include condensing at least a portion of the refrigerant using the condenser heat exchanger.

The method may comprise passing the cooled high pressure refrigerant from the condenser heat exchanger to the first tube bank. The method may comprise using the first tube bank to provide a degree of subcooling to at least a portion of the high pressure refrigerant. The method may comprise providing subcooling of up to 20° C., e.g. between 0° C. and 20° C., to the high pressure refrigerant using the first tube bank. This subcooling may be achieved through heat exchange between the high pressure refrigerant and ambient air and/or through heat exchange between the high pressure refrigerant and the low pressure refrigerant flowing through the second tube bank.

Reducing the pressure of the cooled high pressure refrigerant may comprise providing the cooled high pressure refrigerant to an expansion device after it has been passed through the first tube bank. The expansion device may be an electronic expansion valve. The method may include expanding the refrigerant using the expansion device. The low pressure refrigerant may be passed to the second tube bank, and flow through the second tube bank after it has passed through the expansion device.

The method may include providing a degree of superheat to at least a portion of the low pressure refrigerant using the second tube bank. The method may include providing up to 20° C., e.g. between 1° C. and 20° C., of superheat to the low pressure refrigerant using the second tube bank. This may be achieved through heat transfer from the high pressure refrigerant in the first tube bank to the low pressure refrigerant in the second tube bank and/or through heat transfer from the ambient air to the low pressure refrigerant in the second tube bank.

The refrigerant may be R-32 refrigerant or propane.

The method may include evaporating at least a portion of the low pressure refrigerant using the second tube bank. The method may include flowing the low pressure refrigerant through the second tube bank in a counter-current flow to the high pressure refrigerant flowing through the first tube bank.

Passing high pressure refrigerant through the first tube bank may cause all of, or at least a portion of, the external surface of the first tube bank to be heated to a temperature of at least 50° C. In particular, the external temperature of a row of the first tube bank that is closest to the second tube bank may be at least 50° C.

Passing high pressure refrigerant through the first tube bank will preferably cause heat transfer from the first tube bank to the second tube bank. This may result in the temperature of the external surface of the second tube bank being raised above 0° C. In particular, this may result in the temperature of the external surface of a row of the second tube bank that is furthest from the first tube bank being greater than 0° C. It will be appreciated that, in use, the external surface temperature of the rows of the second tube bank may decrease the further away the row is from the first tube bank. However, the external surface temperature of all of the rows of the second tube bank may be sufficiently high to prevent and/or limit ice build-up thereon.

The method may include flowing refrigerant sequentially through one or more rows of the first and/or second tube bank. The method may include flowing refrigerant through one or more first and/or second tubes that are connected in series within each of the one or more rows of the first and/or second tube bank. The tubes may be configured in a serpentine shape, such that the refrigerant is flown through the row(s) of the first and/or second tube bank in a serpentine shape.

The method may include dividing the flow of refrigerant entering the first tube bank into multiple flows, e.g. two, three or more flows, using a first refrigerant distributor such that the refrigerant flows into one or more refrigerant circuits of the first tube bank. The method may include combining the multiple flows of refrigerant exiting the refrigerant circuits of the first tube bank using a first refrigerant header.

The method may include dividing the flow of refrigerant entering the second tube bank into multiple flows, e.g. two, three or more flows, using a second refrigerant distributor such that the refrigerant flows into one or more refrigerant circuits of the second tube bank. The method may include combining the multiple flows of refrigerant exiting the refrigerant circuits of the second tube bank using a second refrigerant header.

The method may include flowing air over the first and/or second tube bank, such that heat is exchanged between the air and the first and/or second tube bank. It will be appreciated that through this interaction, heat will also be exchanged between the air and the refrigerant flowing through the first and second tube banks. The method may include flowing air over the first and/or second tube bank using a fan. The method may include causing the air to flow over the first tube bank before the air flows over the second tube bank, or vice versa. If the air is first flown over the first tube bank before it flows over the second tube bank, heat from the high pressure refrigerant in the first tube bank may be transferred to the air, thereby heating the air before the air flows over the second tube bank. As a result, the air reaching the second tube bank may be of a higher temperature compared to the ambient temperature of the air (i.e. the air that has not been heated through interaction with the first tube bank). The refrigerant flowing through the second tube bank may therefore be heated to a greater extent through heat exchange interaction with the warmed air flow. Moreover, the warmed air may be of a sufficiently high temperature that the resulting cooling of the warmed air through a heat exchange interaction between the air and the refrigerant in the second tube bank may not lead to the formation of ice on the second tube bank through cooling and freezing of moisture within the air.

The method may include flowing air over the second tube bank in a counter-current flow to the refrigerant flowing through the rows of the first and/or second tube banks.

According to a third aspect of the invention, a method of manufacturing an air-to-water heat pump system may be provided. The method may include connecting, in series: a compressor for increasing the pressure of a refrigerant, a condenser heat exchanger for receiving a high pressure refrigerant output from the compressor and for transferring heat from the high pressure refrigerant to a source of water, a first tube bank of an evaporator heat exchanger, the evaporator heat exchanger being for extracting heat from ambient air, an expansion device for reducing the pressure of a refrigerant received from the first tube bank of the evaporator heat exchanger, and a second tube bank of the evaporator heat exchanger. The first and second tube banks of the evaporator heat exchanger are arranged in close proximity to one another such that, in use, heat from a high pressure refrigerant passing through the first tube bank will be transferred to the second tube bank in order to limit and/or prevent ice build-up on an external surface of the second tube bank.

The method may include connecting an outlet of the second tube bank to the compressor.

The method may provide the air-to-water heat pump system of the first aspect. Accordingly, the method may include arranging the first tube bank and the second tube bank such that a gap is formed between the first and second tube banks. The size of the gap may be 1 mm or less, preferably 0.5 mm or less. The first tube bank may be arranged so as to be in direct contact with the second tube bank at points along its length, or along its entire length. The method may include arranging at least a portion of a first plurality of fins of the first tube bank in close proximity to at least a portion of a second plurality of fins of the second tube bank. The method may include arranging all or a portion of the first plurality of fins such that they are 1 mm or less from a corresponding fin of the second plurality of fins.

BRIEF DESCRIPTION OF THE DRAWINGS

Certain embodiments of the disclosure will now be described by way of example only and with reference to the accompanying drawings in which:

FIG. 1 is a schematic diagram of an air-to-water heat pump system;

FIG. 2 is a schematic diagram of an alternative air-to-water heat pump system;

FIG. 3 is a schematic diagram of an evaporator heat exchanger of the air-to-water heat pump system of FIG. 2; and

FIG. 4 is a schematic diagram of a row of an evaporator heat exchanger tube bank.

FIG. 5 is schematic diagram of a first tube bank and a second tube bank of an evaporator heat exchanger.

DETAILED DESCRIPTION OF THE INVENTION

As shown in FIG. 1, an air-to-water heat pump system 20 comprises a compressor 22, a condenser heat exchanger 24, an evaporator heat exchanger 26, and an expansion device 28 arranged in a fluid circuit to permit a flow of refrigerant through the system 20. The heat pump system may comprise a fan 30, which may be configured to create a flow of air A over the evaporator heat exchanger 26. Typically, in air-to-water heat pump systems such as the heat pump system 20, the evaporator heat exchanger 26 is located outdoors such that it can extract heat from the ambient air. The heat pump system 20 may further comprise a controller (not shown), which may be configured to control certain aspects of the heat pump system 20. For example, the controller may be configured to control the compressor speed and/or the opening degree of the expansion device 28.

During operation of the heat pump system 20, the compressor 22, which may be a scroll compressor, compresses a refrigerant to produce a high pressure refrigerant, which may be partly or completely in gaseous form. The high pressure refrigerant exits the compressor 22 and enters the condenser heat exchanger 24, where it exchanges heat with a flow of water 32. As the high pressure refrigerant has a higher temperature than the flow of water 32, heat is transferred from the high pressure refrigerant to the flow of water 32. Thus, the flow of water 32 is heated whilst the high pressure refrigerant is cooled. The heated flow of water 32 may travel to an indoor area, for example to a radiator, and may provide heating to that indoor area. The condenser heat exchanger 24 may be configured to condense at least a portion of the high pressure refrigerant, converting that portion from the gaseous phase to the liquid phase. However, a portion of the refrigerant may remain in the gaseous phase upon exiting the condenser heat exchanger 24.

After exiting the condenser heat exchanger 24, the high pressure refrigerant passes into the evaporator heat exchanger 26. Specifically, the high pressure refrigerant passes into a first tube bank 34 of the evaporator heat exchanger 26. Here, the high pressure refrigerant may exchange heat with the ambient air, such that the high pressure refrigerant is cooled by the ambient air, and the ambient air is heated by the high pressure refrigerant. Preferably, the refrigerant is cooled sufficiently such that it is all, or at least predominantly, in liquid form as it exits the first tube bank 34. The high pressure refrigerant exits the first tube bank 34 and enters the expansion device 28, where it is reduced in pressure, resulting in the lowering of the temperature of the refrigerant. In the expansion device 28, the refrigerant becomes a low pressure refrigerant.

The low pressure refrigerant exits the expansion device 28 and enters a second tube bank 36 of the evaporator heat exchanger 26. Here, the ambient air may exchange heat with the cooled, low pressure refrigerant, such that the low pressure refrigerant is heated and the ambient air is cooled. This may result in at least a portion of the low pressure refrigerant in the second tube bank 36 being evaporated, causing that portion to change phase from a liquid refrigerant to a gaseous refrigerant. The refrigerant may then exit the second tube bank 36 and be passed back to the compressor 22, beginning the cycle again.

When the ambient air temperature is low, there is a risk of ice forming and building up on the evaporator heat exchanger 26. This may occur when water vapour present in the ambient air encounters the cold outer surface of the evaporator heat exchanger 26 (that has been cooled by the presence of the low pressure refrigerant flowing therethrough), causing the water vapour to condense and, if the air is cooled to below the freezing point of the water (e.g. 0° C.), freeze. As a result, ice may begin to build-up on large portions, if not all, of the outer surface of the evaporator heat exchanger 26. This may prevent the ambient air from directly contacting the evaporator heat exchanger 26, thus limiting heat exchange between the ambient air and the low pressure refrigerant flowing within the second tube bank 36. An excessive build-up of ice on the evaporator heat exchanger 24 may therefore reduce the efficiency of the heat pump system 20, and/or reduce its coefficient of performance (COP). The build-up of ice on the heat exchanger 26 can be a major problem in countries with colder climates, or during the winter months, as outdoor temperatures are frequently sufficiently low that further cooling of the air through interaction with the evaporator heat exchanger 26 causes ice formation on the evaporator heat exchanger 26. In typical systems, ice build-up may begin to occur when the ambient air temperature is lower than 10° C., and the rate of ice build-up will generally increase as the ambient air temperature reduces further.

To counter this issue, heat pump systems frequently have external heaters arranged in close proximity to the evaporator heat exchanger. These heaters may be turned on at given intervals, or may be turned on in response to an indication of excessive ice build-up. This use of external heaters not only requires additional energy in order to power the heater, but also increases the footprint of the heat pump system. Such bulky systems are generally undesirable. Alternatively, a “defrost” mode may be employed by the system when ice build-up is excessive. A “defrost” mode is typically when the flow of refrigerant through the heat pump system is reversed, causing the evaporator heat exchanger and the condenser heat exchanger to switch functions. That is, the flow of refrigerant through the system is typically reversed so that a relatively warm, high pressure refrigerant is flown through the evaporator heat exchanger and a relatively cool, low pressure refrigerant is flown through the condenser heat exchanger. In this way, heat is emitted from the high pressure refrigerant within the evaporator heat exchanger, thus melting ice that may have built-up on the surface of the heat exchanger. However, a lot of energy is required for such a defrost mode, and the use of such a mode disrupts the normal heating function of the heat pump system.

As can be seen from FIG. 1, the first tube bank 34 of the evaporator heat exchanger 26, which is on the high pressure side of the heat pump system 20, and the second tube bank 36 of the evaporator heat exchanger 26, which is on the low pressure side of the heat pump system 20, are in close proximity to one another. This allows heat to be exchanged between the two tube banks 34, 36. During normal operation, the first tube bank 34 will contain a high pressure refrigerant that is at a higher temperature than the low pressure refrigerant present in the second tube bank 36. The temperature of the high pressure refrigerant may be such that the temperature of the outer surface of the first tube bank 34 may be, for example, 50° C. or greater. The outer surface of the second tube bank 34 will typically be much cooler due to the presence of the relatively cold, low pressure refrigerant therein. As such, heat is transferred from the first tube bank 34 to the second tube bank 36. This transfer of heat may occur through conduction (due to the close proximity of the tube banks 34, 36). Additionally or alternatively, heat may be transferred between the two tube banks 34, 36 via convection or radiation. The first tube bank 34 therefore heats an outer surface of the second tube bank 34, for example such that the temperature of the outer surface of the second tube bank 36 is greater than 0° C. As a result, ice build-up on the outer surface of the second tube bank 34 is limited and/or prevented.

As will be appreciated, this arrangement offers various advantages over prior heat pump systems. Evaporator heat exchangers typically have multiple tube banks. By connecting one of these tube banks to the high pressure side of the heat pump system, as in the present invention, the heat from the high pressure refrigerant can be effectively utilised to prevent and/or limit ice build-up on the evaporator heat exchanger without the need to introduce an external source of heat. The footprint of the heat pump system also remains the same. That is, the build-up of ice can be limited and/or prevented without the need to increase the footprint of the evaporator heat exchanger. Additionally, it may no longer be necessary to employ a “defrost” mode as excessive ice build-up can be avoided, or at least the frequency at which a “defrost” mode needs to be used can be significantly reduced, as the rate of ice accumulation can be decreased. As a result, the efficiency of the heat pump system can be increased by up to 20% as compared with prior art systems.

As mentioned above, prior art heat pump systems typically begin to encounter ice build-up when the ambient air temperature is approximately 10° C. or lower. The minimum ambient temperature at which ice will not form on an evaporator heat exchanger is known as the “freeze limit”, and for typical systems this is around 10° C. In the heat pump system 20 shown in FIG. 1, the freeze limit may be reduced significantly as compared to prior art systems. For example, the freeze limit of the heat pump system 20 may be lower than 10° C., and preferably lower than 7° C. The heat pump system 20 may therefore not accumulate ice unless the temperature is lower than 7° C. Further, the rate of ice build-up on the evaporator heat exchanger 26 at a given temperature may also be greatly reduced, even at ambient air temperatures below the freeze limit, as compared to prior art systems, meaning that excessive ice build-up is less likely, and the need for a “defrost” mode is eliminated or reduced.

The heat pump system 20 may be a reversible heat pump system. In other words, it may be possible to reverse the flow of refrigerant in the heat pump system 20 depending on whether a user wants a cooling mode or heating mode. In FIG. 1, the heat pump system 20 is shown in a heating mode, with the flow of refrigerant through the system being shown by the arrows in the refrigerant flow path. In this mode, the refrigerant in the evaporator heat exchanger 26 extracts heat from ambient air, which is then expelled and provided to the water source in the condenser heat exchanger 24, thereby heating the water. If the flow were reversed, the evaporator heat exchanger 26 would function as a condenser heat exchanger and vice versa, thus causing the system 20 to extract heat from the water (e.g. from inside a building) and expel heat to the ambient air outdoors. Since, in this cooling mode, the evaporator heat exchanger 26 is heated by the refrigerant flowing therethrough, the cooling mode may be utilised as a “defrost” mode when necessary in order to melt ice that has built up on the evaporator heat exchanger 26 during operation of the system in the heating mode. A valve, such as a four-way valve, may be utilised in order to reverse the flow, though this is not shown in FIG. 1.

Typically, reversible systems require a receiver to be located in the refrigerant flowpath between the condenser heat exchanger and the evaporator heat exchanger. This can be required because the volume of condenser heat exchangers is typically much smaller than the volume of evaporator heat exchangers, which can result in excess liquid refrigerant within the system when the system is switched from the heating mode to the cooling mode, and this is stored in the receiver before (in a defrost mode) it is passed to the condenser heat exchanger. However, the heat pump system 20 may not require a separate receiver, as the first tube bank 34 may function as a receiver to retain excess liquid refrigerant during the cooling mode. As such, the need for a receiver may be eliminated, further simplifying the system and reducing its footprint.

In FIG. 1, the fan 30 is arranged to cause an air flow A to pass firstly over the second tube bank 36, and then over the first tube bank 34. This may be achieved by positioning the fan 30 adjacent to the second tube bank 36. However, the fan 30 can alternatively be arranged to create a flow of air A that passes firstly over the first tube bank 34 and then over the second tube bank 36. This may be achieved by positioning the fan adjacent to the first tube bank 34, as shown in FIG. 2. The flow of air A may therefore be heated through interaction with the relatively warm, high pressure refrigerant in the first tube bank 34 before being passed over the second tube bank 36, where it may provide heat to the outer surface of the second tube bank 36 and also provide additional heating for the low pressure refrigerant within the second tube bank 36. By first passing the air over the first tube bank in this way, the temperature of the air that is passed over the second tube bank 36 can be raised above the ambient air temperature, and may be raised above the freeze limit of the system. As a result, subsequent cooling of the air through heat exchange with the low pressure refrigerant in the second tube bank 36 may not be sufficient to cause ice to form on the second tube bank 36. Even in instances where ice does still form on the second tube bank 36, the rate of ice formation may be reduced compared to if air at the temperature of the ambient air (i.e. air that had not been warmed) had been passed over the second tube bank 36. The warmed air may also raise the temperature of the second tube bank 36 as it passes over the second tube bank 36, thus helping to limit and/or prevent ice build-up on the second tube bank 36 and also helping to heat the low pressure refrigerant.

In FIG. 1, the second tube bank 36 is shown as comprising three rows, 36a, 36b, 36c, as will be described in more detail below with reference to FIG. 3. This system 20 is arranged such that, during the heating mode, refrigerant is first passed through row 36a, which is positioned furthest from the first tube bank 34, and is then passed through row 36b and then row 36c in series before exiting the second tube bank 36. It will be appreciated that in this arrangement the refrigerant flows towards the first tube bank 34 as it flows through the rows 36a-c of the second tube bank 36. Hence, the refrigerant flows in generally the same direction as the air flow A over the evaporator heat exchanger 26. This may be termed “co-current” flow. Typically, co-current flow is employed in reversible heat pump systems in order to avoid significant performance loss when the system is operated in a cooling mode. Hence, co-current flow may be used where it is desired to provide a good level of performance when the system is operated in the heating mode as well as when the system is operated in the cooling mode. Co-current flow may be employed more regularly in warmer climates, such as Mediterranean climates, where the ambient air temperature is relatively warm.

An alternative arrangement is shown in FIG. 2, in which the second tube bank 36 is arranged such that during a heating mode refrigerant is passed through the rows 36a-c in a direction generally opposite to the direction of the air flow A. This may be termed “counter-current” flow. Counter-current flow typically leads to optimised heating of the refrigerant in a heating mode by the air that is passed over the evaporator heat exchanger, and therefore optimised performance of the system 20 in a heating mode. This may be particularly beneficial in cooler climates where the ambient air temperature is low, such as in Nordic climates. Counter-current flow typically provides for a greater degree of heating compared to co-current flow.

Whilst the system shown in FIG. 1 is arranged for co-current flow, it will be appreciated that it could alternatively be arranged for counter-current flow. Similarly, the system 20 of FIG. 2 may be arranged for co-current flow, rather than counter-current flow.

The arrangement of the evaporator heat exchanger 26 may provide a certain degree of subcooling to at least a portion of the high pressure refrigerant within the first tube bank 34. For instance, the temperature of the refrigerant may be reduced by up to 20° C. as it is passed through the first tube bank 34. This may be as a result of being cooled through heat exchange with the ambient air and/or the second tube bank 36 (and the refrigerant therein). This subcooling may ensure that a majority, if not all, of the high pressure refrigerant is in liquid form before entering the expansion device 28 and, subsequently, the second tube bank 36 of the evaporator heat exchanger 26. As such, there may be no, or very little, gaseous refrigerant entering the expansion device 28 and/or the second tube bank 36. This may improve the efficiency of the heat pump system 20, as a greater amount of refrigerant may be evaporated in the second tube bank 36 and thus a greater amount of heat may be extracted from the ambient air.

The arrangement of the evaporator heat exchanger 26 may provide a certain degree of superheat to at least a portion of the low pressure refrigerant within the second tube bank 36. For instance, the temperature of the refrigerant may be increased by up to 20° C. as it is passed through the second tube bank 36. This may be as a result of being warmed through heat exchange with the ambient air and/or the first tube bank 34 (and the refrigerant therein). This may ensure that the entirety, or at least a large proportion, of the low pressure refrigerant is evaporated and is therefore in gaseous form as it leaves the second tube bank 36. This may ensure that there is no, or very little, liquid refrigerant remaining in the refrigerant that is passed to the compressor 22. This may be important to ensure that the compressor 22 continues to operate correctly, as the presence of too much liquid refrigerant within the compressor can create faults and/or could damage the compressor 22. Any drops in efficiency that may result from improper function of the compressor may therefore be avoided.

Hence, the evaporator heat exchanger 26 of the heat pump system 20 may be considered to be a combined evaporator, subcooler, and superheater.

FIG. 3 shows a schematic diagram of the evaporator heat exchanger 26 of FIG. 2. The first tube bank 34 is shown on the right of the diagram, and the second tube bank 36 is shown on the left of the diagram. The flow of air A is shown flowing from right to left such that it passes over the first tube bank 34 before it passes over the second tube bank 36, similar to the arrangement shown in FIG. 2. The evaporator heat exchanger 26 may be a round tube plate fin (RTFP) heat exchanger, and optionally a microchannel heat exchanger. The evaporator heat exchanger 26 may therefore comprise a plurality of fins 46, 48, such as plate fins. The plate fins may lie in a plane that is perpendicular to the direction of refrigerant flow through the heat exchanger 26, as shown in FIGS. 3 and 4. The first tube bank 34 may comprise a first plurality of fins 46 and the second tube bank may comprise a second plurality of fins 48. The evaporator heat exchanger 26 may comprise copper tubes and/or aluminium fins.

High pressure refrigerant output from the condenser heat exchanger 24 enters the first tube bank 34 through a first inlet 38. The high pressure refrigerant then travels through the first tube bank 34 and exchanges heat with the second tube bank 36. The high pressure refrigerant may also exchange heat with the air flow A. In the first tube bank 34, the high pressure refrigerant is cooled as it expels heat. After passing through the length of the first tube bank 34, the high pressure refrigerant exits through a first outlet 40 and is passed to the expansion device 28 (not shown in FIG. 3). Although shown with only one row in FIG. 3, it will be appreciated that the first tube bank 34 may comprise multiple rows, similar to the second tube bank 36 shown in FIG. 3.

Low pressure refrigerant output from the expansion device 28 enters the second tube bank 36 through a second inlet 42. The second tube bank 36 is shown in FIG. 3 as comprising three rows 36a, 36b, 36c; however, it will be appreciated that the second tube bank 36 may comprise more or fewer rows than this. The low pressure refrigerant enters the first row 36a and then passes sequentially from the first row 36a to the second row 36b and finally to the third row 36c. In FIG. 3, the final row 36c is shown adjacent to the first tube bank 34; however, the first row 36a may instead be adjacent to the first tube bank 34, with the final row 36c being furthest from the first tube bank 34. The low pressure refrigerant then exits the second tube bank 36 via second outlet 44, and is passed to the compressor 22. The rows 36a-c of the second tube bank 36 are heated by the first tube bank 34. As a result, as it travels through the rows 36a-c of the second tube bank 36, the low pressure refrigerant is heated, and at least a portion may be evaporated. The low pressure refrigerant may also be heated via heat exchange with the ambient air, such as air flow A passing over the second tube bank 36. The rows 36a-c may be arranged in parallel with one another and may be arranged in parallel with the rows of the first tube bank 34. As such, a large surface area of each tube bank 34, 36 may be in close contact with one another.

As will be appreciated, the temperature of the row 36a will be increased through the exchange of heat from the first tube bank 34 the least amount by virtue of it being furthest from the first tube bank 34. However, the outer surface temperature of the first row 36a may still be increased to higher than 0° C. due to the heating caused by the presence of the relatively high temperature first tube bank 34 and the air flow A. In this way, it may be ensured that the build-up of ice is prevented and/or limited over the entire surface of the second tube bank 36.

The system 20 may be configured such that the row of the first tube bank 34 that is closest to the second tube bank 36 has a surface temperature of at least 50° C. This has been found to provide adequate heating to the second tube bank 36 to prevent and/or limit ice build-up.

In order to provide for heat exchange between the first and second tube banks 34, 36, they should be in close proximity to each other. As can be seen from FIG. 3, a gap may exist between the first tube bank 34 and the second tube bank 36. This gap may have a size D. Though shown as a gap with constant size in FIG. 3, the gap may vary in size along the length of the tube banks. For example, the gap may have a size of 0 mm along at least a portion of the length of the tube banks, such that the tube banks are in direct contact for at least a portion of their length. Regardless, the gap may have a size D that is no larger than 1 mm (or preferably no larger than 0.5 mm) at any point along the tube banks. As shown in FIG. 3, the gap D may be measured between the extremities of the first plurality of fins 46 and the second plurality of fins 48. This closeness in proximity may aid heat transfer between the two tube banks 34, 36. In some cases, the gap may be sized such that water droplets, e.g. condensed from the ambient air, may bridge the gap between the first and second tube banks 34, 36 to enable heat to be transferred between the first and second tube banks via the water droplet(s).

FIG. 4 shows an example of a single row of a first and/or second tube bank 34, 36. FIG. 3 may be considered a top view of one of the rows 36a-c shown in FIG. 4.

As can be seen, the row may comprise one or more tubes 50 arranged in parallel. Two or more of these tubes 50 may be connected to one another in series to form a refrigerant circuit 52. The tubes 50 may be connected to form, for example, a serpentine shaped or a coil shaped refrigerant circuit 52. A single row of the tube bank 34, 36 may comprise one or more refrigerant circuits 52. In the example shown in FIG. 4, the tube bank 34, 36 comprises three pairs of tubes 50, i.e. there are six tubes 50 in total. Each pair of tubes 50 is connected in series to form a serpentine-shaped refrigerant circuit 52. Hence, the row includes three serpentine-shaped circuits 52. Whilst FIG. 4 shows the specific example of six tubes 50 and three serpentine-shaped refrigerant circuits 52, it will be appreciated that any number of tubes 50 and refrigerant circuits 52 is possible, and any shape of refrigerant circuit 52 is possible. A plurality of fins 46, 48 may extend between the tubes 50 in order to aid heat transfer.

The row shown in FIG. 4 is arranged for use in a tube bank 34, 36 that includes only a single row, and has a refrigerant distributor 54 for receiving refrigerant fed into tube bank 34, 36 via an inlet 38, 42. The refrigerant distributor 54 acts to divide up the flow of refrigerant into different refrigerant streams and separate the refrigerant between each of the refrigerant circuits 52. After having travelled through the tubes 50 of a circuit 52, the refrigerant may then be re-combined in a refrigerant header 56 before it exits through the outlet 40, 44.

Either or both of the first tube bank 34 or the second tube bank 36 may comprise one or more rows. If a tube bank 34, 36 includes more than one row, the refrigerant may pass through all rows before it is re-combined. In this way, a refrigerant circuit 52 may span across more than one and/or all of the rows in a tube bank 34, 36. It will therefore be appreciated that not each row in the tube bank 34, 36 will require a distributor 54 and/or a header 56. Rather, only the first row (i.e. having the inlet 38, 42) may include a distributor 54 and only the last row (i.e. having the outlet 40, 44) may include a header 56. The use of one or more refrigerant circuits 52 may provide redundancy to the system. The use of the one or more refrigerant circuits 52 may also lead to lower pressure drops in the refrigerant, and increase the thermal efficiency of the tube bank 34, 36. In this way, the heat exchange performance of the evaporator heat exchanger 26 may be optimised.

FIG. 5 shows a cross schematic view of an evaporator heat exchanger 26 comprising a first tube bank 34 with a single row 34a and a second tube bank 36 with a plurality of, in this case three, rows 36a-c. In this Figure, the layout of the refrigerant circuits 52 can be seen.

In FIG. 5, the first tube bank 34 includes a first refrigerant distributor 54a into which high pressure refrigerant may enter via the first inlet 38. The first distributor 54a acts to divide the flow of refrigerant and pass it to a plurality of high pressure refrigerant circuits 52a, which each comprise a plurality of, in this case two, high pressure refrigerant tubes 50a fluidly connected in series. After passing through the high pressure refrigerant circuits 52a, the refrigerant from each of the high pressure refrigerant circuits 52a is passed to a first refrigerant header 56a, where the refrigerant is recombined into a single flow before exiting the first tube bank through the first outlet 40. In FIG. 5, the first tube bank 34 comprises a single row 34a comprising twelve refrigerant circuits 52a that each comprise two high pressure refrigerant tubes 50a. However, it is envisioned that the first tube bank 34 may comprise any number of rows 58a with any number of refrigerant circuits 52a.

The second tube bank 36 includes a second refrigerant distributor 54b in which low pressure refrigerant may enter via the second inlet 42. The second distributor 54b acts to divide the flow of refrigerant and pass it to a plurality of low pressure circuits 52b. As shown in FIG. 5, each of these low pressure refrigerant circuits 52b may span across multiple rows 36a-c of the second tube bank 36. Each row 36a-c comprises multiple low pressure refrigerant tubes 50b. Each low pressure refrigerant circuit 52b may comprise a plurality of these low pressure refrigerant tubes 50b connected in series. FIG. 5 shows a second tube bank 36 that comprises twelve refrigerant circuits 52b that extend across three rows 36a-c. In the illustrated example, each circuit 52b includes two low pressure tubes 50b in the first row 36a, two low pressure tubes 50b in the second row 36b, and two low pressure tubes 50b in the third row 36c. However, it will be appreciated that the second tube bank 36 may comprise any suitable number of refrigerant circuits 52b, comprising any number of tubes 50b and extending over any suitable number of rows 36a-c. After being passed through the low pressure refrigerant circuits 50b, the refrigerant is passed to a second refrigerant header 56b. The recombined low pressure refrigerant may then exit the second tube bank 36 through the second outlet 40.

The numbers of tubes 50, circuits 52, and rows 60 may be chosen based on a number of factors, including (but not limited to) performance target, heat exchanger footprint, refrigerant properties, and fin density.

As the two refrigerant flows pass through the first tube bank 34 and the second tube bank 36 respectively, they exchange heat with one another, thus heating the exterior surface of second tube bank 36 such that ice build-up on the exterior surface of the second tube bank 36 is prevented and/or reduced.

The evaporator heat exchanger 24 of the present invention prevents and/or eliminates ice build-up by utilising heat from the high pressure refrigerant in the first tube bank 34 to maintain and/or increase a temperature of the second tube bank 36. This evaporator heat exchanger 24 can be achieved by simply modifying an existing heat exchanger such that one or more of its rows may be utilised as a first tube bank (for a high pressure refrigerant) and one or more of its rows may be utilised as a second tube bank (for a low pressure refrigerant). This offers advantages in terms of space, complexity, and cost, as well as leading to improvements in efficiency for the heat pump system 20. A simple and efficient heat pump system 20 is therefore provided by the use of such an evaporator heat exchanger 24.

Claims

1. An air-to-water heat pump system comprising:

a compressor for increasing the pressure of a refrigerant;

a condenser heat exchanger arranged downstream of the compressor for receiving a high pressure refrigerant output from the compressor and for transferring heat from the high pressure refrigerant to a source of water;

an expansion device for reducing the pressure of a refrigerant; and

an evaporator heat exchanger for extracting heat from ambient air, the evaporator heat exchanger comprising:

a first tube bank having a first inlet arranged to receive a high pressure refrigerant output from the condenser, and a first outlet fluidly coupled to the expansion device; and

a second tube bank having a second inlet arranged to receive a low pressure refrigerant output from the expansion device, and a second outlet fluidly coupled to an inlet of the compressor;

wherein the first and second tube banks are arranged in close proximity to one another such that, in use, heat from the high pressure refrigerant passing through the first tube bank is transferred to the second tube bank in order to limit and/or prevent ice build-up on an external surface of the second tube bank.

2. The air-to-water heat pump system as claimed in claim 1, wherein a gap exists between the first and second tube banks, the gap being 1 mm or less in size, preferably 0.5 mm or less in size.

3. The air-to-water heat pump system as claimed in claim 1, wherein the first tube bank is arranged in a heat exchange relationship with ambient air such that, in use, heat is transferred from the high pressure refrigerant to the ambient air, thereby cooling the high pressure refrigerant and warming the ambient air; and/or

wherein the second tube bank is arranged in a heat exchange relationship with ambient air such that, in use, heat is transferred from the ambient air to the low pressure refrigerant, thereby warming the low pressure refrigerant and cooling the ambient air.

4. The air-to-water heat pump system as claimed in claim 1, wherein the first and/or second tube bank comprises one or more rows arranged in parallel with one another and fluidly coupled to one another in series.

5. The air-to-water heat pump system as claimed in claim 5, wherein each row of the first and/or second tube bank comprises a plurality of first and/or second tubes connected in series with one another, optionally wherein the plurality of first and/or second tubes are arranged in a serpentine or coil shape.

6. The air-to-water heat pump system as claimed in claim 1, wherein the first and/or second tube banks comprise a plurality of refrigerant circuits.

7. The air-to-water heat pump system as claimed in claim 1, wherein the first tube bank comprises a first plurality of fins and the second tube bank comprises a second plurality of fins, the first plurality of fins and the second plurality of fins being in close proximity to one another.

8. The air-to-water heat pump system as claimed in claim 1, wherein the system is configured such that the freeze limit of the system is less than or equal to 10° C., preferably less than or equal to 7° C., the freeze limit being the minimum ambient temperature at which ice will not form on the heat exchanger.

9. The air-to-water heat pump system as claimed in claim 1, wherein the system is configured such that, in use, the temperature of the external surface of the first tube bank is at least 10° C. higher than the temperature of the external surface of the second tube bank.

10. The air-to-water heat pump system as claimed in claim 1, comprising a fan arranged to create a flow of air over the evaporator heat exchanger, preferably wherein the fan is configured to flow air over the first tube bank before the air flows over the second tube bank.

11. A method of preventing and/or limiting the build-up of ice on an evaporator heat exchanger in situ in an air-to-water heat pump, the method comprising:

passing a high pressure refrigerant through a first tube bank of the evaporator heat exchanger in order to cool the high pressure refrigerant though heat exchange between the high pressure refrigerant and ambient air;

reducing the pressure of the cooled high pressure refrigerant, so as to provide a low pressure refrigerant; and

passing the low pressure refrigerant through a second tube bank of the evaporator heat exchanger in order to warm the low pressure refrigerant through heat exchange between the low pressure refrigerant and ambient air;

wherein the first and second tube banks are arranged in close proximity to one another such that heat from the high pressure refrigerant passing through the first tube bank is transferred to the second tube bank in order to heat at least a portion of an external surface of the second tube bank, thereby limiting and/or preventing ice from building up on the external surface of the second tube bank.

12. The method as claimed in claim 11 comprising:

increasing the pressure of the refrigerant with a compressor, thereby providing the high pressure refrigerant, before passing the refrigerant to a condenser heat exchanger; and

prior to passing the refrigerant to the first tube bank, using the condenser heat exchanger to cool the high pressure refrigerant through heat exchange between the high pressure refrigerant and a water source, thereby warming the water source.

13. The method as claimed in claim 12, drawing air over the first and second tube banks, preferably wherein the air is drawn over the first tube bank before flowing over the second tube bank.

14. The method as claimed in claim 12, wherein the refrigerant is R-32 refrigerant or propane.

15. A method of manufacturing an air-to-water heat pump system, comprising connecting, in series:

a compressor for increasing the pressure of a refrigerant,

a condenser heat exchanger for receiving a high pressure refrigerant output from the compressor and for transferring heat from the high pressure refrigerant to a source of water,

a first tube bank of an evaporator heat exchanger, the evaporator heat exchanger being for extracting heat from ambient air,

an expansion device for reducing the pressure of a refrigerant received from the first tube bank of the evaporator heat exchanger, and

a second tube bank of the evaporator heat exchanger,

wherein the first and second tube banks of the evaporator heat exchanger are arranged in close proximity to one another such that, in use, heat from a high pressure refrigerant passing through the first tube bank will be transferred to the second tube bank in order to limit and/or prevent ice build-up on an external surface of the second tube bank.